Cross drive for heavy vehicles



Nov. 28, 1961 0. K. KELLEY 3,010,342

CROSS DRIVE FOR HEAVY VEHICLES Original Filed July 21. 1947 10Sheets-Sheet 1 attorney O- K. KELLEY CROSS DRIVE FOR HEAVY VEHICLES Nov.28, 1961 Original Filed July 21. 1947 8 g I I 1 Nov. 28, 1961 0. K.KELLEY 3,010,342

CROSS DRIVE FOR HEAVY VEHICLES Original Filed July 21. 1947 10Sheets-Sheet 3 UIIIIIIIIIIIIA WIIIIIIIII Snnentor Gttornegs Nov. 28,1961 0. K. KELLEY 3,010,342

CROSS DRIVE FOR HEAVY VEHICLES Original Filed July 21, 1947 10Sheets-Sheet 4 Ihwentor (Ittornegi Nov- 28, 1 0. K. KELLEY 3,010,342

CROSS DRIVE FOR HEAVY VEHICLES Original Filed July 21, 194'? 10Sheets-Sheet 5 .W 9g 11/ W W w w I \E t/flfl /I! 21'; w: "$127 Z- a! Ia; .5 )7 2 Z/. I

1 5 z/a' o a; Zr" a Ciltorneg Nov.. 28, 1961 0. K. KELLEY CROSS DRIVEFOR HEAVY VEHICLES Original Filed July 21. 1947 10 Sheets-Sheet 6 Nov.28, 1961 0. K. KELLEY CROSS DRIVE FOR HEAVY VEHICLES 1O Sheets-Sheet 7Original Filed July 21. 1947 Bnventor (Ittorncg Nov. 28, 1961 0. K.KELLEY CROSS DRIVE FOR HEAVY VEHICLES l0 Sheets-Sheet 8 Original FiledJuly 21, 1947 31mm tor Q r K 6 NW 0. K. KELLEY 3,010,342

CROSS DRIVE FOR HEAVY VEHICLES l0 Sheets-Sheet 9 Nov. 28, 1961 OriginalFiled July 21, 1947 Gttornegs 0. K. KELLEY CROSS DRIVE FOR HEAVYVEHICLES Nov. 28, 1961 Originl Filed July 21. 1947 l0 SheetsSheet 1OGttornegs United States Patent 24 Claims. (Cl. 74-710.5)

This is a division of my Patent 2,596,931, filed July 21, 1947, andissued May 13, 1952, and a continuation of abandoned divisionalapplication S.N. 58,196.

The present invention relates to drive mechanism for large heavyvehicles such as military tanks, tractors and the like, having tracklaying mechanism steered by variable speed ratio difierential means.

It relates more particularly to such mechanisms in.

which quick steering and reversing of direction of travel is required,and for which the drive mechanism embodies an arrangement of unitsproducing plural and continuously variable torque components combinedadvantageously for the stated purposes.

A primary object is to provide a common drive to right and left handtruck drivers which includes final drive output gear units commonlydriven thru divided torque paths, one of which is supplied by coupleddifierential gearing driven from a prime mover and the other of which isdriven by combined selective change speed gearing and fluid torqueconverter mechanism, for the purpose of obtaining maximum smoothness oftorque during speed ratio transition periods and for obtaining similarsmoothness in steering effect created by establishing of a variablerange of selected reaction torques in said differential gearing.

The present invention representing improvements over the inventiondescribed in my Patent No. 2,585,790, filed April 16, 1945 and issuedFeb. 12, 1952, in certain important particulars, it is obvious that thegeneral objects stated in that application are likewise sought herewithinsofar as the subject matter of both applications would permit.

A further object is to provide a power transmission system for tracklaying vehicles having concentric primary and final drive mechanismswith parallel shaft differential mechanism coupling to the said finaldrive mechanisms so as to provide divided and recombined torque for thepurposes above stated, of the same hand of rotation of said mechanisms,for forward drive, the advantages including improved power output withless wear due to lower friction losses.

An important object is the provision of a high degree of compactness ofthe driving and steering assembly obtained by nesting of the variablespeed drive, the differential unit, and the output unit parts, resultingin extremely low torsional couples in the supporting structures withbalancing out of the reaction couples ordinarily experienced in devicesof this character.

In the disclosure herewith another important object is achieved, that ofproviding exceptionally useful efliciency in the variation of thesplit-torque ratio favoring the transfer of torque to the fluid torqueconverter portion of the drive with increased converter eificiency,resulting in an extension of useful torque converter speed range havingas a further resultant a rising engine speed characteristic in thisprocess.

An additional object is to provide a dynamic steeringdriving assemblyhaving continuously variable speed ratio drive compounded from dividedtorque paths one'of which includes a fluid torque converter adapted todeliver a variable torque component to each of the vehicle 3,010,342Patented Nov. 28 1961 tread-driving, final drive output gear units, andthe other chanically connected and adapted to deliver additivecomponents of torque to said output gear units, a supplementary objectbeing to provide in this assembly the desirable characteristics that thedivision of torque between the paths varies to favor the assumption of ahigher torque with increase of efliciency of the fluid torque converter.7

' A further object is the provision, in such an assembly as statedabove, of means for varying the vehicle steeringradius to increase anddecrease with vehicle speed for sharp steering at low speed andstabilized steering at high speed, and of means for pivoting the vehicleat standstill by the same means utilized for said variable radiussteering.

It is an important object herein to provide a driving and steeringassembly which shall utilize the power dividing and recombining featureoutlined above and which shall apply the variable steering torquecomponent delivered to the final drive output units thru a mechanicalgear train or trains while applying the continuously vari-' able torqueconverter and connecting reduction gear trains, the ratios of which areselective for plural forward, and reverse drives. I

It is an added object to provide herein a dynamic power steered anddriven assembly for vehicles in which under given selected driveconditions, may embody means for reapplying a torque from the output tothe input FIG. 1 is a transverse sectional view of the driving assembly,showing the input shaft from the engine geared I to the primary powerdrive, and the output sprocket drivers at the right and left. 7

FIG. '2 is a detailed section view of the difierential couplingarrangement of FIG. 1 for the powered steering effect derived thru thecompensator group at the top of FIG. 1. FIG 3 is a part section taken at33 of FIG. 2 to show the relationships of the diflerential gearelements.

FIG. 4 is a section taken at 44 of FIG. 1 to show the operatingmechanism for the reaction brake 45 for low gear drive. FIG. 5 is asimilar part section taken at 55 of FIG. 1 to show the somewhatdifferent arrangement for the actuation of brake 46 which controls thespeed of the output shaft on the right of the assembly of FIG. 1. Itshould be noted that brake 51 is controlled in accordance with FIG. 4,and brake 50 in the manner of FIG. 5. FIG. 6 is a sectional detail ofthe external mechat the right of FIG. 1, modified by the substitution ofa' disc brake for the output sprocket shafts of FIG. 1.

FIGS. 9 and 10 show elevation sections in right angle planesrespectively for the mechanical connections to the control valves forthe steering action applied by fluid pressure to the differentialsteering clutches of FIG. 1.

FIGS. 11 and 12 represent the drive-selection valve and its externalcontrol, respectively for the modification sys tem shown in FIG. 14 indiagrammatic form. The two steering valves of FIGS. 9 and 10 are shownin the upper right corner of FIG. 14 and the ratio selection valve of 3FIG. 11 is shown adjacent to them toward the center of the diagram. FIG.13 is a sectional detailof a vacuumresponsive valve adapted to cushionthe ratio down-shift effectby' controlling the application of linepressure admitted to the transmission drive system'by'the shifter valveof FIG. 11. The diagram of FIG. 14 shows the valve of FIG. 13 at theright of the shifter valve.

FIG. 14 is a schematic diagram of a fluid pressure system for controland operation of the drive structures of FIG. 1. FIG. 15 is adetailedsectional view of the pressure, regulator valve ofFIG. 14. FIG. 16 is asectional detailed view of auxiliary valving energized during the powersteering interval. as controlled by the FIG. 14 arrangement. i a

j FIG. 17 is aschematic diagram of a combined fluid pressure andelectricalcontrol system, as a modification of the FIG. 14 system,wherein the ratio-controlling valving is by individual valves,electrically operated, and other diiferentiations are provided, asexplained further in detail. FIG. 18 is a view, similar to that of FIG.15,

of the pressure regulator valve of FIG. 17. FIG; 19 is a detailedsectional view of one of the solenoid-actuated control valves shown inFIG. 17, wherein thecontrol impulses such as supplied from an operatorscontrol sys-' tem, as shown in FIG. 20 diagrammatically, are convertedto fluid pressure flow response. 7 ,FIG. 21 is a diagrammatic view ofaform of pump whichmay be used to replace certain of the pumps of FIGURES14 or 17, as required. I

In my above mentioned Patent 2,585,790 the output units consisted ofvariable speed input sun gears and output shaft connected carriershaving power steering drive applied to annulus gears. V p a The presentapplication has output units operating in the same general manner exceptthat the variable speed ratio input: is applied to the annulus gears andthe power steering differential drive to the sun gears. V

In both, the split torque principle is used, of dividing the inputtorquethru two trains, variable. speed drive and-power steering, andcombining, the torques inthe output gear units.

-' This dividing and combining principle is shown in a somewhatdifferent form in my Letters Patent U.S.v

2,176,138 issued October 17, 1939, in the drive for the so-calledHydromatic transmission, and in my Letters Patent US. 2,211,233 issuedAugust 13, 1940.

In the present arrangement, the differential power train consistsofshafts 36 and 55 geared at 41, 25 and 42, 39 to output sun gears 24 and38, the shafts being normally coupled by clutch C for. unitary rotationin the same hand of rotation, this featurebeing different from thearrangement in Serial No. 588,475, where .the secondary power group hastwo shafts geared for oppositerotation.

For economy of space and for resolution of drag couples to oneconcentric alignment, the power steering differential centerline isherein concentric with the output unit centerline. 1 p a g It has beenfound that the introduction of the splittorque principle into thecompound fluid turbine and gear drive assemblies embodyingdynamic-steering provides a highv attainable efiiciency over comparativeseriesunit drive assemblies without impairing the desirable drivecharacteristics of the fluid units. r

In the present invention the power, is deliveredthru a mechanical pathand a hydraulic and methanical path, the first including a planetarygear train having planet. gear carrier rotating fixedly with the outputshaft, and having a sun; gear and ring gear one of whichcouples beexplained in detail further;

4 gear assembly D which provides steering eflect but has no resultantaction directly upon the aforesaid torque dividing operation. 7

The annulus gears 18 and 33 are cross-connected thru drum flanges ofcentral shaft'40 and are driven by the output member of the fluid torqueconverter unit W. For neutral drive, this connection is interrupted aswill The construction shown in FIG. 1 has engine shaft 1 furnishinginput power, geared at 3, 4. to hollowinput shaft 5. The two outputsprockets or load members 21 and 37 are each driven from adjacent finaldrive planetary gear units. The shaft delivers power to the carriermembers 6 of the centrally located diiferential gear unit left by a drum17 to annulus gears 16 and 18 of the reversing and final drive unitsadjacent shaft 21.

I The combining output gear group 33, 34, 38 at the right, drivescarrier 35 and output sprocket 37 as a. final drive unit, whilecombining output group 18, 19, 24 at the left drives carrier 20 andsprocket 21 similarly.

The reduction gear group 29, 27, 12' provides low gear and coupleddirect drive to shaft 40 from shaft 7, and the reverse gear group 16,13, 12 provides reverse gear drive to shaft 40 from shaft 7. V

The gear groups 1819--24 and 33'*34 38 should be regarded astorque-combining variable ratio gears, and the shaft 40with drums 17 and28 may be thought of as the cross-coupling input means for thetorque-combining gears, while the combination 29 27.12 is the lowreduction group and the combination 16'13-12 is the reverse reductiongroup. The clutch 3fi --31 drives the coupling shaft 40 at unit speed.with shaft 7, when energised.

The line of drive from the unit W therefore consists of final torquecombining groups driven by change speed groups. Since the member .40 isthe means for providing equalcouplingto the annulus gears18, 33 of eachof ordinate distribution of the torques.

The output sprocket drive members 21 and 37 are connected to and drivenby the output unit carriers 20 and 35 The sun gears 24 and 38fof theseunits aredn'ven from the power steering'diiferential unit D, and theannulus gears 18 and 33 from the fluid torque converter W and connectedgear train.

In straight, non-steering drive, the secondary shafts 36, 55 geared todrive the output unit sun gears are coupled for unit rotation by thefriction clutch C, and the powersteering differential annulus gears4949' rotate at the samespeed and in the same direction, causing theplanet to'the' output of the fluid torque converter unit W, the

drive members 21 and37, and the sun gears 24*and38: I

are connected to the input side of the fluid torque converter W thrugears 25, 39, 41 and 42 and a differential pinions 52 to stand still,and their carrier,6 to couple the turbine output member 0 to .bothannulus gears 49, 49' at unit speed and rotation.

The'engine-connected shaft 1, bevel geared at 4'drives the input shaft 5which rotates the turbine impeller I and the power-steering differentialcarrier 6. vThis divides the engine torque thru the turbine W and thepower steering differential train coupled to the sun gears 24, -38 ofthe output units. V V a The intervening solid shaft 40 cross-connectingthe output units is attached thru drums 17 and 28 to the annulus gears33, 16 of both output units. With one component of torque applied thruthe secondary shafts 36, 55 from the steering differential D to the sungears 24, 38 of the output units, and another applied from the fluidturbine W and gearing to the annulus gears 13, 33 thereof, the carriers20, 35 of the output units will revolve at a differential ratio of thecomponents received, which will vary in accordance with the load andspeed conditions, and drive unit characteristics.

The drive is initiated by locking of low gear brake 45 to stop rotationof annulus gear 29. Assuming that the converter W may deliver a torqueto sun gear 12 this torque is applied to carrier drum 28 of shaft 40, onthe one hand, and is also applied thru shaft 7 to sun gear 12, theannulus gear 16 of drum 17 being driven. The component of this torque isapplied to drums 17 and 28 of shaft 40, and consequently to annulusgears 18 and 33. Simultaneously shaft 5 is driving carrier 6, planetpinions 52 and annulus gears 49, 49' of the power steering differentialD at input speed, and the teeth 48 and 48' of the mnulus gears arerotating gears 43 and 44, drums 55' and 57, shafts 55, 36 and gears 41,42, the latter elements revolving in the same direction at the samespeed.

Consequently output sun gears 24 and 38 deliver the same torque fractionto the output units, and the output carriers 20 and 35 receive equalcombined torques.

For straight-ahead running, shafts 55 and 36 rotate together at the samespeed, assuming the tractive efforts 'on sprocket drive members 21 and37 are the same, since clutch C compels them to do so. While it may seenpossible to omit cluch C, on the assumption that variation is right andleft drive tractive effort could be compensated for by the difierentialaction in unit D, it must be remembered that over uneven ground, if thearrival of one side at a point where traction is less permitsdifierentiation, the passing of the same side to firmer ground couldinroduce a yawing tendency opposite in steering effect to that firstexperienced. The clutch C therefore couples the power steering train andits path of torque evenly to both output units, while the other paththru the torque converter W and its gearing trains is likewise evenlycoupled.

At above a desired given speed, the clutch 30-31 is locked, and theturbine output shaft 7 rotates the shaft 40 and annulus gears 13 and 33at turbine output speed.

The recombining of divided torque in the output units is readilyunderstood, as resulting from additive or subtractive components createdin each unit by sun gear and annulus torque producing a resultantsummation torque on the output carriers 20 or '35 and members 21 or 37.

For all practical purposes shaft 4% and drums 17 and 28 serve as thefirst train input members for the torque combining groups for each trackdriver.

The variable speed gear assembly of the first train consists of torqueconverter W receiving power from shaft 5 and gear 4, and delivering sameto the cross-coupling shaft 40 acting as the output element of theseparated forward and reverse reduction groups driven by shaft 5 thrusun gears 12 and 12.

The one-way clutch F coupling shaft 5 to sun gear 12, acts to by-passthe torque converter W and couple shaft 7 to shaft 5 when there isvehicle motion or rotation of shaft 7 with no motion, or a lesser speedof shaft 5. This assures that shaft 7 will not ever exceed the speed ofshaft 5, and enables the operator of a vehicle to obtain a towed startof a stalled engine, for example.

While it would be possible to place clutch F so as to couple shaft 4% toshaft 5 on the overrun at l-to-l ratio, it must be remembered that shflt40 is required to be driven reversely, so that lockout means for clutchF would be needed, to permit this reverse rotation of 40 to 5. Thepositioning shown in the figures herewith is simple and requires noauxiliary lockout since the point of torque 6 conversion for reversedrive lies beyond the intermediate coupling of shafts 5 and 7 by theclutch F.

While it is appreciated that the prior art shows one-way clutchesarranged to by-pass variable speed ratio drives and to couple output toinput at l-to-l ratio on the overrun, the problem herein solved is toutilize this effect for by-passing a torque converter combination whichmay have low overtaking torque efficiency, and doing so in theintermediate connection to the step ratio variable speed gearing of thetrain, in order to retain steering stability control on downhill runs.

The overall driving train herein shown provides a powerful low speedratio reduction range, and since such drives under overtaking torqueendeavor to speed up input driving elements to extremely highoverspeeds, the drive system must be protected against the possibilityof unrestrained run-away, as well.

Assuming the installation to be in a military tank and running on a downgrade at considerable speed, if the fluid of the torque converter W weresuddenly drained as may occur in battle from an enemy shot, the momentumof the vehicle would no longer be restrained by the reverse drag of thetorque converter W, and the reverse step-up in ratio thru the gear unitsto shaft 5 would not be loaded or restrained by engine braking. Clutch Ftherefore acts to couple the converter-connected shafts at 1 to 1 beforedamaging high speeds under reverse torque are reached.

If such a mishap occurs when one of the reaction brakes is energised,instead of at a time when clutch :3631 is engaged, the spinning speedsof the planet gears of the assembly could reach damaging velocities, andthe same would be true of a mishap to the controls for the low andreverse brake bands 45 or 51, respectively.

The brakes 50 and 46 for the respective output carriers 20 and 35 of thetorque combining units enable the operator to stop the motion of eitherof sprockets 21 or 37, together or selectively.

Carrier 20 being stopped, for example, by brake 50, the rotationalcomponent applied to drum 17 by shaft 40' is transferred thru planets 19as a reverse component to sun gear 24 and gear 25, the latter rotatinggear 41 shaft 55 in the same hand of rotation as drum 17. This rotationis transferred to gear 44, which rotates crown gear 49 and reacting thrugear 52, speeds up gear 49.

Assuming normal forward drive, the carrier 6 of the differential unit Dwould rotate clockwise as viewed from the left, but gear 49 would be inopposite rotation. Clutch C is disengaged for this action.

Since the carrier 6 has forward clockwise motion and the gear 49 hasreverse rotation thereto the gear 49' revolves with a resultant motionat increased speed in the same hand of motion as shaft 5. Thistransmitted thru gear 43, shaft 36, gear 42, to sun gear 38 adds adifferential component to the coupling of the combining unit 33-3438greater than was obtained thru normal operation, so that the sprocketdrive member 37 may be rotated faster.

.This action causes the vehicle to pivot or steer about a point adjacentthe stopped member 21, and causes the other tread to advance faster.

In reviewing this peculiar relationship of elements, one should considermember 21 and carrier 20 of the output unit at the left of FIG. Istopped by brake 56, while the carrier acts as a reactor, so that torquemay be transferred between sun gear 24 and annulus 18. If the gear unitat the right is in low gear, and the steering brakes 50, 51 areinactive, shaft 7 drives the annulus gear 33 at a reduction consistingof the variable ratio of the unit W multiplied by that of the right-handgear unit, while the normal component derived thru the rotation of theunit -D is capable of being transferred to shaft 36, gears 42, 39 andsun gear 38. However, since shaft 40 is driving the other output unitannulus 18 at the same speed as annulus 33, the sun gear 24 instead ofhaving a fixed ratio forward component derived from D, is now urgedbackward at an overspeed ratio by rotation of annulus 1S. 7

' Assuming this torque increment is transferrable to unit D, thedifferential transfer gear 48 would endeavor to rotate backward whilethe normal rotation of unit D would be urging'the piece 47, 48, 49 torotate forward. The clutch C'being locked or engaged, no difierentialmotion in unit D could take place. If clutch C has sufficient brakingcapacity, the reaction of the couple would then have the effect ofbraking drum 17, shaft 40, drum 28 and hence rotation of shaft 7, sothat the net result of applying only one output shaft brake such as 54?would be to brake the whole forward motion of the vehicle, withoutsteering effect; 7

Now if clutch C be released, and brake 50 beapplied, the backwardrotation of member 47,748, 49 would be transmitted across pinions of thedifferential unit D, and member 47', 48', 49 would tend to rotateforward faster than differential carrier 6 and shaft 5. This speeds upthe steering torque delivery shaft 36, gears 42., 39 and sun gear 38 ofthe left-hand-output unit, so that a sharpsteering effect isimmediately-obtained, by reason of the stop pingof the member 21 whichestablishes a pivot for the vehicle, for the fulcrurning action of theincreased speed of carrier 35 and member 37. a

It should be remembered, in studying this mechanism that the normalslowing of member 47, 48', 49, for example, by steering device B, alsoslows the rotation of gear 43, shaft 36, gears 42 and 39 and sun gear38, tending to cause the vehicle to steer'toward that side of thevehicle.

The odd dual effect obtained by applying one side brake, like 50, tostop one output element directly while accelerating or speeding up thedrive to the other side output element, by merely opening across-connecting clutch, is believed entirely new in this art.

' The power steering path of torque begins with the shaft drivingcarrier 6 of the differential unit D. The gears'49, 49 are meshed withthe planet gears 52 and have external teeth 48, 48 meshing withdifferential output gears 43, 44. The divided shaft 36, 55 turns thegears 41, 42in the same hand of rotation, but of opposite hand to therotation of shaft 5, and of gears 25, driving the sun gears 24, 38 ofthe output units.

. The drums 47, 47 of crown gears 49', 49 extend radially and laterallywhere steering brake effects areapplied to graduate the retardationrequired to obtain the desirable steering effects.

FIGS-2 and 3 may be consulted at this point to clarify the action of thedifferential mechanism. The steering devices A and B for the compositemembers 47, 48, 49 and 47', 48', 49' are shown in part detail in FIG. 2,the plates 61 'being splined to rotate with part 47, and the plates 62being attached to part 64? of the casing. \The prime member, parts atthe right of FIG. 2 are similarly attached. 'Disregarding for the momentthe power steering effect and assumingthat the shafts 36 and 55 aredriven together at unit speed from the engine by shaft 5, differentialcarrier 6, crown gears 49, 49' and coupling gears 43, 44-,the fluidtorque converter output member 0 may drive hollow shaft 7, and assumingthat the sprocket-shaft carriers 2! 35 are under equal torque andtractive load, the coupling pattern of the gearing requires that wheneither of bands 45 or 51 are energised to hold their drums 29 or 15, thereaction established by annulus gear 29 or carrier 15 compels the inputpower to be expressed as driving torque.

With band 45 locked for low gear drive, the carrier 28' and shaft 40 aredriven in the same hand of rotation by sun gear 12 at a reduction ratio,while in the other output unit drum 17 and annulus 18 are rotatedequally and similarly since they are connected by shaft 40.

The torque converter W may then be operated over its useful torquemultiplying range, multiplied further by the reduction ratio of theoutput units. If the input sun gearsI-Z and 12! of the output unitscould be held against rotation, the annulus gears 18 and 33 would drivethe carriers 35 and 26 at a fixed reduction. However, the power steeringtrain is providing a fixed ratio drive from engine connected shaft 5',in the same hand of rotation, as will be obvious from a tracing of therotations of the train elements. 7

Assuming that the differential unit D transmits power equally to the sungears 24 and 38 of the two output members 21 and 37, and since theannulus gears 18 and 33 are connected by shaft 40, the drive operationin low gear proceeds by clamping brake band 45 on drum 29 of annulusgear 2 which causes carrier 28 and attached annulus. gear 33 to rotateat a reduction ratio, which rotation would be transmitted to carrier 35of output member 37 by whatever reaction or rotational component isapplied to sun gear 38 from the mechanical train composed of the shaft5, differential group 6, 49, 48, 43, shaft 36, gear 42 and gear 39driving the sun gear 38.

The engine power applied to hollow shaft 5 divides one component passingthru the torque converter W to shaft 7, sun gears 12 and 12', finaldrive annulus gears 18 and 33; the other component passing thru thedifferential group D to shafts 36, 55, gears 41, 42 and gears 25 and 39coupled to the final drive sun gears 24 and 38.

The application of clutch,3t)31 with release of band 45, couples annulusgear 29 with sun gear 12, compelling shafts 7 and 40 to rotate together,which places a similar couple across annulus 16 and sun gear 12 of thereverse gear group, so that the speed of fluid torque converter' outputmember 0 is applied to annulus gears 18 and 33 equally, the sun gears 24and 38 receiving their components from the diflerential group D, shaft55, 36, gears 41, 42 and 25, which latter drive the sun gears equally.

Alternate actuation'and operation of clutch 36-31, with brake band 45provides two ranges of forward speed ratios, entirely suflicient for allthe needs of a large heavy vehicle.

To enable one to visualize clearly the relative rotational components ofthe various elements involved in this drive, the FIGURE 1 should bere-examined for the fact (that the combined torque components forforward drive in the final output gearing are both of the samerotational hand as that of input; thatis, the hands of rotationof shafts5, 7, 40, of sun gears 24 and 38 and of output carriers 20 and 35 arethe same.

The known prior art does not explain the utilization of the dividingtorque principle in differential power drives, not does a generalknowledge of this appear to be known in engineering practice or textbooks, hence it is believed proper herein to dwell at some length on it,that the invention may be properly and clearly understood. 7

Referring back to the conditions under which divided torque is obtainedin low gear, it may be stated that the output unit sun gear 24, and ringgear 18 which participate in the recombing of the torque should haveequal tooth loading under all drive conditions, and that the variablespeed ratio torque path delivers a multiplication of approximately 3.5to 1, in the present example, as shown, to the ring gear 18.

For ease of calculation, the tooth load times the pitch radiusrepresents torque. The ratio of the gears 18 and '24 is 2.5 to 1,therefore annulus gear 18 is always loaded 2.5 times more than sun gear24. The speed ratio of the differential drive mechanical connectionbetween sun gear 24 and the converter input shaft 5 is about 7 to 1,therefore the incoming torque is that of the sun gear divided by 7,while that of the annulus gear is 3.5 times greater because of thetorque multiplying planetary gear between the converter output member 0and the annulus 18.

The torque converter output and input torques are related by a factor Rwhich varies with output speed, there- 9 fore the annulus gear torquedepends on this factor, and upon converter output speed.

For clarity this statement may be set up as an equation: Converter inputtorque on shaft 5 sun gear torqu EQUALS The mechanical fraction of theinput torque is:

sun gear torque 3.5 converter R PLUS 7 Since sun gear torque divided by7 equals the mechanical torque, one can write: Engine torque EQUALSMechanical torque 3.5 converter R PLUS 1) and therefore the expressionfor the percentage of input power carried by the differential mechanicalpath is: Mechanical torque Engine torque 3.5 converter ratio R EQUALS17.5 PLUS 3.5 converter ratio R It should be understood that thisdisclosure is based on the mechanical dimensions and factors and only byway of example, and only serves to show the steps of reasoning by whichthe useful results are achieved.

Upon being given the torque multiplication ratios of the torqueconverter for the different output speeds, one may thereupon by theabove process determine the percent of the divided torque going thru thetwo paths.

The overall efl'iciency obtained on a speed chart is a flatter curvethan with other comparative drives, and the corresponding engine speedcharacteristic has an increasing upward slope. The percentage of torquetaken by the mechanical drive, in the example herewith given will varybetween 15 and 50 percent, actually experienced.

When the clutch 30-31, is engaged for the high forward drive, the 3.5 to1 ratio train between converter output shaft 7 and the final drive unitsat the left and right of FIG. 1 is locked in direct drive. By thisshifting or changing of the relative speed range, the percentage ofdivided torque is transferred to a difierent scale in which thepercentage of the torque thru the mechanical train will vary between 22and 5 percent.

This relationship of the divided and recombined torque factors isoperative only in the two forward speed ratios.

When reverse band 51 is applied to carrier drum 15 for planet gears 13,the divided torque factors are reversed, and instead of the input powerbeing divided into two paths, the arrangement produces the unusualeffect of increasing the drive input torque to a higher value than thatbeing developed by the engine.

An analogy of this would be in a self-energized brake mechanism in whichthe brake anchor or reaction force may be re-applied thru linkage toincrease the primary energizing forces.

In FIG. 1, the band 51 holding drum 15, the reverse gear group 161312applies a reverse rotational component to annulus gear 18. If member 21is not rotating, the sun gear 24 would rotate forward at 2.5 times thespeed of the gear 18, which thru the 7 to 1 ratio of the gearing train25, 41, 44 rotates the torque converter 10 input shaft 5 at 17.5 timesthe speed of the annulus gear 18, or 7 times faster than the speed ofconverter output shaft 7.

Therefore, with the vehicle not in motion, the torque converter outputshaft 7 is rotating forward at oneseventh of input speed and the energyrepresented by this rotation of shaft 7 is fed back to the torqueconverter input shaft 5 through the same gearing which in forward speeddrive provided a recombining of torques.

This novel result is not mere text-book theory, but represents actualmechanical results proven by test instrumentation, as Well as bysuccessful operation of large vehicles on the road. v

This exact percentage of feed back or return flow of torque to input maybe calculated:

2.5 torque converter output torque 2.5 X 7 EQUALS torque converte;output torque Since the converter speed ratio is 7 to 1, the torqueratio is determinable from experience tables, which show that it isapproximately 3.8 to 1, thereforetorque converter output torque EQUALS3.8 torque converter input torque and it follows that:

torque converter input torque MINUS torque converter feed back torqueEQUALS engine torque Since torque converter output torque divided by 7equals the feed back torque, it follows that:

7 times feed back torque EQUALS 3.8 times torque converter input torqueor expressed another way:

torque converter 3 .8

input torque EQUALS torque converter feed back torque torque converterinput torque EQUALS 1.84 times torque converter feed back torque Havingobtained a proportionality for the relationships between torqueconverter input and feed back torques, one may then find out what thefeed back value is, in terms of initial engine torque.

Having above, that:

times torque converter input torque MINUS torque converter feed backtorque EQUALS engine torquewe can substitute:

1.84 times torque converter feed back torque MINUS torque converter feedback torque EQUALS engine torque therefore .84 times torque converterfeed back torque EQUALS engine torque and rewriting this:

torque converter feed back torque EQUALS 1.19 times engine torque Thisodd effect therefore results in inducing an in- 11 and left tread forcereactions are equal, and exactly divided by the differential action.

The steering brakes are not energized for' straight drivi-ng. Clutch5456 being normally engaged, shafts 36 and 55 deliver equal torques tothe final drive output gear units, hence the output shaft speeds to thetread sprockets are exactly equal to each other, and equal in' if athigh converter speeds, the resulting turning radius is expanded so thatproportionally steadier steering is provided for high speed travel asagainst steering at low speeds by quick maneuvering with a short turningradius.

This effect of continuously variable steering radius increasing withdrive speed is believed of exceptional novelty and utility in providingvery accurate steering for large heavy vehicles under all drive surfaceand gradient conditions.

Common forms of slipping brake controls in this field of art arenotorious for high energy dissipation requirements, and in the presentinvention, the matching of the drive characteristics of the unitsdescribed herein produces a. locked drum steering sequence which avoidsneed for excessive brake energy dissipation. It should be noted thatregardless of vehicle speed, the speed efiect derived thru themechanical ditterential drive path is constant with constant enginespeed, while that derived thru the hydraulic drive path is proportionalto vehicle speed, since at constant engine speed the only otherpermissible variable in the over-all drive is-the vehicle speed.

A further useful result herein is the fact that when the output members21 and 37 are non-rotating, vehicle stopped, the application of one ofthe steering difierential brakes A or B results in one tread beingdriven forwardly and the other reversely, which causes pivoting of thevehicle on its own center. *In military' practice, this feature is ofexceptional value in enabling the tanks to reverse direction in a narrowspace, and it adds also to operating facility in agricultural and dirtmoving machinery.

FIG. 4 shows a hydraulic actuator for low band 45 which mechanism isalso used for actuating reverse band 51. The band is pivoted at 77 tostrut piece 78 supported in casing 100 by adjustable stud 79 and nut 80.The movable end of the band is pivoted at 76 to strut 75 fitting a notchor recess in lever 73 pivoted to the casing at 74. The lever 73 carriespivot 72 for link 71 of piston rod 70a of piston 70 operating incylinder 68 formed in the casing. Fluid pressure is admitted at 81 tohold or push piston 70 against spring 69, and pressure on the oppositefaceof 7G is admitted to and released from cylinder 63 thru passage 82.p

The valve control system for the hydraulic servo action 7 is shown inFIGS. 9 to 20 inclusive, and described further herein. 7

The brake control of FIGS. 5 and 6 pertains to the operation of thevehicle brakes 4-6 and 59. When these brakes are applied, it isdesirable that cooling lubricant be flowed at considerable velocity overthe braking surfaces.

Band 46 is energized by pivoted strut 83, notched arm 84, link 85,rocker arm 86, shaft 87 and external lever 88 supported in casing 100.

The lever 86 is rocked clockwise to energize brake 46.

12 Main pump line pressure from the oil cooling system is connected topipe 90, and pipe 91 is connected to an opening between the band '46 anddrum 28. The valve plug 92 is recessed to hold pressure in line 90,.andto withhold it from pipe 91, and is held normally in the position shownwhen brake 46 is not applied. Projection 86:: .or arm 86 registers asshown, until lever 88 causes arm 86 to swing for energizing band 46.Pressure in 9%} thereupon moves valve 92 to the right, uncovering theupper end of pipe 91, so thatcooled oil flows to the space betweenband'46 and drum 28; When external lever 33 is moved to release brake46, the projection 36a forces-valve 92 to the left, covering the endof'pipe 91, stopping the fiow from pipe 90. This device provides brakecooling when it is needed, during the actuation period, and shuts itoil? when the cooling capacity'is required in some other group of thesystem.

PEG. 7 shows the divided shafts 55 and 36 coupled by clutch C in asomewhat different space relationship of the par-ts of FIG. 1. In FIG. 7shaft 55' is splined to drum 53 with which plates 54 rotate and externaldrum 57 splined to shaft 36 carries plates 56 mating with plates 54.

Th clutch C of FIGS. 1 or 7 is for the purpose of maintaining anequivalent torque on both output sprocket drive members at all timesexcept when a steering action is desired.

In FIG. --7 is shown the clutch assembly with plates 54 and 56 connectedto drums 53 and 57 of shafts 55 and 36 respectively. Cylinder 64 isconnected with passage in shaft 55 as indicated by the broken lines inthis figure. The piston63 in cylinder 64- is moved. by pressure inpassage 65 against a set of common disengaging springs, not shown. Theclutch arrangement is concentric with shafts 55 and 36. The fluidpressure in tpassage 65 is controlled by valve 220, FIG. 16 which trolis described further in detail below, in connection with FIG. 16 and in-a further system'modification, in FIG. 17. 7

FIG. 7 shows a bevel gear differential system in place of the spur gearunit of FIGS. 1 to 3, and also shows a modified steering brake discussedin detail below in connection with FIGS. 9 and 10, pertaining tosteering method. The drums 47 and 47' are splined to the axial sleevesof difierential members 48 and 48' which are in turn toothed externallyto mesh with gears 44 and 43.

a FIG. 8 is a detailed view of the right hand portion of FIG. 1,modified by the substitution ofa disc brake construction for the bandbrakes of FIG. 1 for the vehicle. ..In FIG. 8 modification, a frictiondisc brake 102, 10s is substituted for both [the band brakes 46 and 50of FIGS. 1 and 5,other novelsfeatures appearing. The output turbine O ofthe converter W is connected to shaft 7 thru splining, which shaft isshown integral with sun gear 12 and clutch drum 30a for plates 30;Planet gears 27' meshing with 12 mesh externally with annulus gear 29and are supported on pins 27 of carrier 28', fixed to annulus gear '33meshing with planets 34. Planets 34, supported on pins 34' of carrier28, mesh internally with sun gear 38'attached to gear 39. Carrier 28 isattached to sprocket wheel shaft 37 to drive same. Drum 29' of annulusgear 29 supports plates 31 mating with plates 30 and is formed to makeacylinder 217 for presser piston 215. The external rim of drum 29"is'surrounded by brake band 45 as in FIG. 4, which is actuated by themechanism described in connection with FIGS. 4 to 6.

In FIG. 8 the drum 28 is splined externally to carry matching teeth offriction 'discs 102 mating with discs 103 splined to fitting 104attached to the casing 100, a

13 tached to overhanging member 105 having a radial flange for takingendwise thrust from springs 106 seated in a recess of fitting 104. Ring107 is anchored in channel 108 between the overhang of fitting 104 andmember 105', and may transmit thrust to the adjacent loose ring 109 thrua ring of bearing balls 109.

The second ring 109 lies in the channel between fitting 104 and member105, to the right of ring 107, and is fitted with a short arm 99 rockedexternally by appropriate controls.

The facing radial surfaces of rings 107 and 109 are machined intoovoidal section recesses, not numbered, in which balls 109 are mounted,so that when relative rotation between the rings occur, the balls 109'ride into the shallower portions of the ovoidal seotion recesses, andexert an axial thrust tending to press ring 109 to the right to shiftfitting 105 to the right against springs 106 thereby shifting plate 105'to clamp the plates 102 and 103 together. This form of loading mechanismprovides a high degree of mechanical advantage for energizing thebraking action for stopping shaft 37' and sprocket drive member '37, inobtaining a pivotal steering action for the dual tread vehicle, or foractual braking of the vehicles drive motion.

The clutch plates 30, 31 of FIGS. 1 and 8 are released by springs, asshown, and are clamped by piston 215 against the opposing flange of thedrum 29 of annulus gear 29. Piston 215 is recessed in cylinder 217formed in the drum 29 and better shown in FIG. 8. The fluid pressurefeed to cylinder 217 is from passage 218 (FIG. 8) in the casing ofconverter W to a groove and passage 219 in the sleeve of shaft 7 leadingto a groove registering with passage 214 in the sleeve of drum 29', thepassage 218 being shown in full line in FIG. 1 and in dashed line inFIG. 8. The pressure control system covering the pressure feed to thelow and high range drive servo cylinders of FIGS. 1, 4 and 8 is shownfurther in connection with FIG. 14.

Lubricating fluid from the pumping system is fed by connecting passage252, FIG. 1, to central passage 238 in shaft 40, continuous withpassages 258' in carrier shaft 37 as shown in FIG. 8. Radial passage292, axial passage 293 and diagonal passage 294 feed lubricant to. thegear group 122729, and radial passage 295 delivers it to the bearings ofplanet gear 27. Similarly, an unnumbered passage delivers oil frompassage 258' to the spindle 34 of planet gears 34. Spent oil flows backto the sump as understood further in examination of FIG. 14.

Referring back to FIGS. 2 and 3, passage 330 in shaft 40 may becontinuous with passage 258 of FIG. 8, and feed lubricant thruintersecting radial passages 331, grooves 332, and passages 333 in shaft7 to channels 334 formed in the inner radial portion of carrier 6 forthe difierential gearing lubrication.

The general lubrication system is amplified in the discussion of FIGS.14 and 17.

FIGS. 9 and 10 show the mechanism for operating the differentialsteering valves 170 and 171 to control the feed to the steering brakeactuators of FIGS. 1, 2 and 14.

In the preceding drive description the mechanical steering efieetsobtained by graduated braking of diiferential members 47 and 47' weredescribed. Fluid pressure cylinders 58 and 58' formed in thenon-rotating members 60 and 60' of FIGS. 1 and 2 enclose annular pistons59 and 59 respectively. Feed passages 172 and 173 of FIG. 1 connect thecylinders 58 and 8' to the ports 174 and 175 of the steering valves ofFIG. 10.

In FIG. the valves .170, 171 he in parallel bores 176 and 177, eachvalve having two end bosses e and f and protruding stems 178 and 179.

Mounted on the stems 178 and 179 are slidable spring seat collars 181and 182 which are secured on these stems by unnumbered lock rings.

.Shaft 180 located and supported in valve body 200 at right angles tothe plane of the valve bore centers, is fitted with a rocker plate 184having fingers 186 and 187 at either end registering with a channel inthe slidable spring seat collars 181 and 182.

The external arm 188 is held on shaft 180 by a serrated section and nut,and carries roller 189 for coaction with external mechanical controls.Seal plugs 191 permit removal for valve repair or adjustment.

Each bore 176v and 177, is equipped with three ports, in order from leftto right, exhaust, clutch feed and main line delivery. The middle port174 for valve connects to the cylinder 58 of steering brake A forvariable braking of member .7 of FIG. 1, whereas the middle port forvalve 171 connects to the cylinder 58' of steering brake B. Each valveis centrally drilled from the blind end of boss 1 at the right to apoint to the left of that boss, and the narrower neck pontion iscrossdrilled at 194 and 195 to connect the spaces between the bosses tothe bore and spaces 196 and 197.

In the vertical position of rocker plate 134 both exhaust ports 198 and199 are connected to ports 174 and 175 open to the actuator pressurefeed passages 210 and 211 and leading to cylinders 58 and 58 as in FIG.2 and the bosses f seal the pressure inlet ports 201 and 202.

Rocking of the rocker plate 184 clockwise tends to shift boss 2 of valve171 to the left to open the exhaust port 199 wider, While valve 170 ismoved to the right to close exhaust port 198 and open pressure deliveryport 201 to thesteering brake feed port 174 and to line 210, of FIG. 14for clamping plates 61, 62 of FIG. 2 to slow down rotation ofdifferential member 47 of FIG. 2, for the desired steering efiect.

The valve action will be described for valve 170, in which the linearspacing between bosses e and f is so taken with respect to the spacingof the ports 198 and 201 that a very close control over the pressureacting in the cylinder 58 of the piston 59 being actuated, isobtainable. The cross-drilling 194 connected to end space 196 has theeffect of metering a proportional fraction of the line pressure to thespace 1% behind the valve, and since the interior pressure between thebosses is equalized to prevent direct axial force on the valve, theend-space pressure over the end face of the valve provides a. forceproportional to valve cross-section area tending to shift the valve awayfrom feed actuation toward the left, in which it will occupy theposition shown in the drawings herein.

In other words the valves 170 and 171 tend to unload me steeringdifierential brakes automatically.

However, the operators control consisting of arm 138 of shaft 180,worked by external rnecham'srn coacting with roller 139 pinned to thearm 188, applies a force acting thru calibrated springs 203, 204, theforce pattern providing a range of equilibrium points in the motioncapable of being felt as a small variable reaction force on theoperators control arm 188, so that in steering the vehicle, the operatoris able to maintain a very smooth, tactile adjustment on the steeringvalve action. It is believed novel to provide pressure reaction feel incontrol devices for manual valves which regulate the diiferentialsteering of drive mechanisms such as described herein.

The resulting self-compensation between spring-force, valve position andnet pressure in end space 196 provides automatic adjustment of the netsteering brake pressure.

It is not deemed necessary to describe the opposite steering action bywhich valve 171 regulates the actuation of differential actuator B formember 47' of FIG. 2. The overall control diagram of FIG. 14 will bebetter understood by reference to FIGS. 9 and 10.

Auxiliary cooling for the steering brakes A and B is provided for, onemethod being shown in which the pressure passages 172 and 173 of FIG. 14are connected to jet passages 322 and 321 respectively; delivering to apair 7 of jets 318, 321 one for each steering brake. This methodprovides additional cooling flow during all intervals when the steeringbrakes A or B are under fluid pressure actuation. A second method inwhich the cooling flow is taken from a lower pressure portion of thesupply system is shown in FIG. 17.

FIG. 17 shows a modified form of steering difierential in which a bevelgear replaces the spur gear difierential of FIG. 2; in which thestraight drive clutch 54, 56 is shown nested inside the compensatorgears 43 and 44 between shafts 36 and 55; and in which the difierentialsteering brake members 47 and 47' are braked by electrical means.

In my Patent 2,585,790 referred to, is shown a schematic electricalsteering control in which appear a simple divided resistance circuitconnecting each resistmce half with a corresponding electrical brakingcoil which receives an increased current with steering angle, from asource of electrical power, as controlled by a selective resistance arm,manually operated. In that disclosure, the energis-. ing current isdelivered thrus a governor-operated switch which cuts off the steeringcurrent at below a given governor speed, but the governor switch may beby-pased by a manual switch if it be desired to utilize the electricalsteering action under speed conditions wherein the gov! ernor would haveinterrupted the power steering circuit. The manually-operated steeringcontactor arm of the Patent No. 2,585,790 disclosure may be used as acut-off switch by being placed in the non-steering middle positiontemporarily, and further, the governor of that showing may be replacedby a simple manual switch. The field magnets are designated by numerals67 and 67 in FIG. 7.

Present reference to the copending application showing is to provideadequate means for steering control by the modification structure ofFIG. 7.

In that figure, the drums 47 and 47' rotating with the members 48 and 48serve the same purpose as drums 47 and 47' of FIG. 2. The field coils 66and 66 respectively supply actuation energy for brakes A and B. Forsteering braking of differential member 48 coil 66 is energized,generating a magnetic flux of proportional value in the plural poleassembly resulting in a braking'of drum 47, since the assembly insidewhich coil 66 is mounted, is bolted to an extension 100d of the casing,as shown. This braking method is old and well-known in the art aseddy-current braking. g This phenomenon is obtained in squirrel-cageinduc: tion or in synchronous motors, designed with salient poles andDC. excitation, the squirrel-cage member being usually a solid iron orsteel drum or cylinder as a rotor in which the eddy currents areinduced. The DC. excitation is used to obtain smooth adjustment of thetorque capacity, which decreases at lower speeds.

Inthe FIG. 7 showing the drums 47 and 47' are the iron or steel rotors,in which the eddy currents are induced, proportional to the currentselectively applied to cores 66 and 66. Since the desired steeringeffect is more gradual than a direct stopping effect for other thanemergency fast pivoting steering, the electrical method is adapted toprovide a fine control of net steering over a considerable and wideturning range. The structural nesting of the differential gearing, thesplit-torque delivery gearing 43, 44 and the steering brakes Aand B ofFIG. 7 demonstrates an advantageous feature of the modification showing.Jet cooling of drums 47 and 47' is described further in detail below.

It is not thought essential to reproduce herein the above-noted FIG. 3of my previously mentioned Patent 2,585,790, for a comprehension of thefull utility of FIG. 7 of the present disclosure. 5

FIG. 11 should be oriented with FIG. 14.

In FIG. 11, valve 110 is shown located in a sectioned portion of valvebody 2439. The external end of the valve the diagram of '110 is moved byarm 111 of shaft 112, thrus pin 113 intersecting two bosses of thevalve, so as to convert rotation of shaft 112 to rectilinear movement.of the valve 10. The FIG. 12 view clarifies this motion coordination.Arm 112 .of shaft 112, with roller 141 coacts with the operatorscontrol. V

For ease in identifying the parts, the bosses of the valve are letteredfrom top to bottom as a, b, c and d. As shown in FIG. 11, the bore 114forvalve 110 in body 200, is intersected by port spaces numbered in thesame sequence 115, 116,117, 118, 119, 120, 121 and 122.

The upper two port spaces 115 and 116 are connected to the pressure feedline 208 and passage 81 of FIG. 14 to deliver fluid pressure to theservo cylinder for the piston operating reverse band 51 of FIG. 1, withan actuator structure equivalent to that of FIG. 4, operated by i 250connected to the pressure delivery sides of the pumps see, 301 of FIG.14. V V V The fourth port 118 is cross-connected by passage 123 in thebody 200 with the seventh port 122 for reasons to be explained furtherbelow. g f i The fifth port 119 is the feed port for the actuation ofband 45 of FIGS. 1 and 4 and connects to the passage 81 of FIG. 4 forthat purpose, thru feed line 261. The'line 261 is connected laterally toby-pass valve 260 of FIG. 13. The sixth port 120, is connected to'exhaust or to the spent-pressure passages leading to'the sump 3-14.

The seventh port 121, is connected to passage 7 259 leading to thepassages 218, 219 and 214 tothe'cylinde'r 217 for actuating thehigh-range clutch piston 215 of FIG. 8; and is connected by passage 166to the chamber 165 at the bottom of the regulator valve bore 153 of FIG.15, to act on the lower face of the bossi, under certain controlcircumstances.

The space 124 at the base of the distributor shifter valve 110 is opento exhaust passage 1%. When the boss a of valve 110 is below the ports1-15 and '116, these ports may drain upwardly past the narrow neck ofthe valve stem, into space 127. v I I g I 4 The ratio-determiningpositions for vtalve .110 are marked on the drawing of FIG; 14, in orderfrom th'e top R, H, L, and N, representing reverse, high range,

low range and neutral, respectively. a

In the next H position the boss b is'stationed between ports 118' and119, so thatline pressure'in port 117 may pass to port 118, thru passage123 to ports 122 and 121 for delivery to the high-range clutchfeedpassages .259 an 218, and boss d seals port 122 from exhaust.

Reverse port 115 at this time is opened to exhaustat 127 and low-rangeport 119 is exposed to exhaust port between bosses b and c. t

The next station downward is at point L'of"the:

sequence marked in FIG. 14, in which the upper edge of boss a is at thelower edge of reverse feed port 116; the

' lower edge of boss b is at the upper edge of line-connected port 117;boss b blocks release from low range port 119 to exhaust port 120, andbosses c and d prevent exhaust from the cross-connected ports 118and122. Feed is therefore from the pumpline port 117 to low-range port 119for actuation of low gear band45 of FIGS. 1 and 4.

The last end-station downward is N, and; registers the lower edge ,ofboss a with the upper edge of port 118; places boss b between exhaustport 120, and high range 17 port 121; places boss bet-weencross-connected port 122 and space 124, and leaves boss d out of the wayin space 124. In this circumstance the main feed port 117 is sealed byboss :1 and all of the feed ports 115, 116, 119 and 121 are open toexhaust.

The valve 110 may be freely moved among the four positions, and in eachposition for delivering servo actuation pressure to the variouscylinders prevents any wrong motion delivery by the peculiar port andboss arrangement, believed to possess points of novelty in this art.

It is not deemed necessary to show the external mechanism motion foroperating valve 110 beyond arm 112' and roller 141 since poppetingactions for valve-stationing and port registry are old in the art. Theroller 141 for example may be traversed over a poppeted sector,the'interpoppet spaces being located angularly to correspond toequivalent angles for arm 111 as determining the stop stations marked inFIG. 14.

Persons skilled in the art may adopt this teaching as desired or needed,without exercise of invention, and obtain the useful result of thepresent device.

It is proper to review FIG. 14 in connection with the by-pass valve 260shown in FIGS. 13 and 14 located in the pressure feed line 261, betweenthe distributor valve 110 and the cylinder 68 of the low range driveactuator piston 70 for brake 45 (FIG. 4).

Engine vacuum derived in passage 262 from the engine intake manifold(not shown) is admitted to space 263 so as to vary the suction pull ondiaphragm 264 acting against spring 265. The valve plug 260 seats at 266against the pressure of line 261 tending to lift the plug 260 from theseat 266 and by-pass fluid into space 267 connected to the spentpressure line 268, which line pressure meets the resistance ofcalibrated spring 265.

The diaphragm 264 is equipped with abutment pin 269 which may strikeadjustable stop pin 270 at a given valve opening spacing.

When pressure is admitted by valve 260 to line 268, the existence of ahigh degree of vacuum tends to draw the diaphragm to the right in FIG.13 permitting the plug 260 to move unrestricted to the right for fullport opening at 266. If the degree of engine vacuum is low as when theengine may be under heavy load, the force of spring 265 tends to limitthe opening between plug 260 and seat 266, so that the amount of fluidrelieved is less and the rise of low range actuating pressure may occurmore rapidly.

The force of spring 265 is chosen to hold the valve closed normallyagainst the pressure of passage 261.

This action is conditioned by the accelerator lever motion and setting.If the drivers lever setting, for example lever 375 of FIG. 20 isdiminished toward idling, the manifold vacuum force may increase,causing the plug 260 to bleed ofi line pressure to a predetermined lowvalue, whereas if the lever is advanced toward a higher engine powersetting, the vacuum force drops in value, and the spring 265 loads plug260 on the seat 266, to close oi the by-pass line 268 thereby causingthe line pressure for establishing the low range actuating pressure toremain high, thus providing high reaction torque capacity.

The initial low pressure phase results in fast low-range actuation underdownshift from relatively low reaction torque capacity for brake 45,whereas the high pressure phase produces a higher torque capacitybuild-up more quickly.

The degree of low-range actuation is therefore commensurate with enginethrottle position, and the invention herein in this particular isbelieved to represent novelty over prior art disclosures in whichvariations in the degree of vacuum are utilized in devices which modifyautomatic ratio changing controls, and in which the force available forclutching is so varied.

The mechanism of FIG. 13 may be built into a common control housing suchas valve body 200, as numbered herein, or placed elsewhere for theconvenience of the operator, as laid out by the designer.

FIGURE 14 is given to instruct the reader'on the'co ordination of themany functions and operations involved in the control, actuation,working fluid supply, lubrication andcooling of the drive mechanism. u

The pressure supply sources are located at the lower right of thefigure, the main line regulation and servo pressure delivery controls atupper center; the steering controls of FIG. 10 upper right, theconverter supply and regulation at the left with the cooling system forthe converter, and the cooling for the friction brake members, at theright center. The FIG. 11 distributor valve appears upper center in FIG.14 and the low-range shift pressure control valve 260 of FIG. 13adjacent, at the right. The

main line regulator valve of FIG. 14 appears in detail in FIG. 15 andthe special control valve 220 for the straight steering clutch C ofFIGS. 1 and 7 is shown in FIG. 16, the control feed line for valve 220being shown at 225.

The preceding text has described the construction andoperation of theshifter valve 110, the special low-range shift control valve 260, thesteering brake valves and 171, and in connection with FIGS. 2, 5 and 7has stated the operation of a portion of the brake cooling system.

In FIG. 14 a plural-pump system supplies four distinct operative groups;(1) the servo force for ratio range shift and for the steering, (2)maintenance of the working fluid of the torque converter under positivepressure, (3) cool-' ing supply for the system fluid, particularly theconverter, the sprocket drive members and steering brakes, and (4)lubrication of all wear surfaces of the entire drive assembly.

Oil is drawn from the sump 314 thru the screens shown at the bottom ofFIG. 14 and is routed thru a special filter 311 locatable in the valvebody 200, which body may include the pressure regulator valve 150' tohold the main line pressure at a high level when gear drive is in thelow-andreverse ranges and at a lower pressure level when drive is in thehigh range.

A second regulator valve 230 in the relief passages from the main lineregulator valve 150 serves to adjust the pressure in the converterworking space feed and cooler lines to a lower than line level. Therelief from valve 150 is the point of take-0E for the generallubrication oil, and this section of the system is held at apredetermined low pressure below that of the preceding section by valve230.

The drain from the converter working space is passed thru two coolers Kand K. Passage 257 from pump 302 is provided leading to cooling jets310, 312 for the vehicle brakes 46 and 50, for constant cooling, whilepassages 91 and 91 provide additional cooling during the brakinginterval as described for the FIG. 5 showing.

Pump 300 driven from shaft 303 geared to shaft 1 of FIG. 1, deliverspressure whenever the engine isoperating. It feeds in two paths, thefirst by line 306 to check valve 305 and to main pressure line 250, thesecond by line 307 to filter 311 and to line 250. If the filler 311becomes blocked, check valve 305 loaded for a given poundage by-passesthe filter 311. Safety valve 308 is connected to relieve excess pressurein line 307 to lubrication line 253 and lubrication feed passage 258.

Pump 301 is driven from gear 299 of FIG. 8 proportionally to outputspeed, and fulfills two needs, one to furnish vehicle brake coolingpressure and the other to increase or augment the pressure of line ofpump 300. The latter result obtains from connecting delivery line 90thru check valve 312' to line 307, with the valve calibrated to provideaugmented feed at a given line pressure. Line 90 feeds as shown in FIG.5 thru valve 92 to cooler jet 399, and thru a second valve 92' to asimilar jet for the other vehicle brake 50.

Pump 302 is also driven from output-speed 299 of FIG. 8 and supplies aconstant cooling feed to the jets 310, 312 and similar jets for brake 50of FIG. 1.

It should be understood that the cooling jet method described here inconnection with the vehicle band brakes 102, 103 to return the sump 314of FIG. 14, as at 317" of FIG. 8.

In the controls of FIG. 14 the valve body 200 may contain onetransmission ratio shifter valve 110 and the two steering brake controlvalves 170 and 171, operated thru external levers 188, 112' shown inFIGS. 9 and 12, these valves directing the flow of oil under pressure tothe servo cylinders for the clutches and brakes. The

' a body 200 may also contain the main regulator valve 150, theconverter regulator valve 230 as well as the other members of the valvecontrol system, appropriately con nected. The steering brake controlvalves 170 and 171 may be controlled by cross lever 184 inside the valvebody 200, which pulls one valve plunger into closed posi tion whiledepressing the other plunger for the other valve, as describedpreceding.

It is not deemed vital to show the complete body 200, since noparticular ingenuity is required to group these elements'with connectedpassages and portings for the carrying out. of the present instructions.

The converter regulator of FIG. 14 consists of tapered valve 230 seatedat 231 in a portion of casing 26%) and normally isolates spaces 232 and232, the space 232 being continuous with the bore 234. The lower'sp ace232' is connected'by angular passage 162 with passage 162 connected toport 161 of valve 159, and also to passage 235 jointed to the twooutflow leads 236 and 236' of the coolers K and K. The upper space 232and bore 234 are connected by passage 237' with the lubrication channel252, with'the line 253 leading fromthe interior of the safety valve 308,with the connection 268 from port 266 of valve 260, of FIG. 13 andlubrication passages 330, 334 and 334 leading to the bevel andtransfergear casings.

The stem 230 of the valve 230 is fitted with a guide piston 229 in aseparate bore portion, and spring 238 loads the tapered portion on-t'neseat 231 at a predetermined force value. Diagonal port 239 drilled inthe tapered head of valve 230 connects space 232 with space 232 evenwhen the valve 230 is firmly seated, so as to deliver lubrication oil tothe connected system.

The space above the piston guide 229 is connected externally at passage233 for drain to the spent pressure passages 259. I

The working space of the converter W as shown in FIG. 14, is fed bypassage 325, the oil being delivered by passages such as indicatedinFIG. 17, adjacent the low pressure Zone of the converter, and thecirculatory flow from the converter passes at 254 to cooler intakes254', from a point adjacent the high pressure zone, as shown in FIG. 8.

This circulation system tends to isolate the body of oil used in theworking space from the body used for lubrication, until the workingspace body has been thoroughly cooled.

It should be noted that cooled oil in cooler outlets 236, 236 mayrecirculate thru space 232' of valve230, angular passage 162 and line252, by thermal and by conespecially pertaining to maintenance of properlubrication. It must be appreciated that in drive mechanism capable ofmoving large vehicles of weights approaching tons, it is of paramountimportance'to control every portion of the drive in which heat isgenerated, not only for the fluid torque converter W described above,but also the other drive elements and units. The cooling lubricationflow for the steering brakes and vehicle brakes is referred to above andin more detail following.

Pressure passages 251, 162, 237 and 252 are involved in the controlledlubrication feed, although the series of exhaust passages numbered 314to correspond to the sump may be tapped, if required to assureadditional oiling of running parts. Line 258 is actually the centralshaft passage of shaft 40 of FIGS. 1, 2 and 8 and feeds lubricant asdescribed above'in connection with FIG. 8.

Line 237 connected to passage 330 of FIG. 1 to furnish lubricant to thediiierential gear of FIG. 2, and is also connected at 330' to passage334' for feeding lubricant to the transfer. gearing of the vehicle, notshown in the drawings. Spent pressure from the transfer gear casing is,of course, drained back to sump 314 of FIG. 14. The valve 257'upper-left in FIG. 14 is set at a given low poundage'to sustain thelubrication pressure in the system between the valve and the lubricationhead pressure furnished in line 251.

The brake and clutch cooling system described in connection with FIGS.2, 5, 7, 8 and 14 is believed to present features of novelty. In largeheavy vehicles of the types for which the present drive assembly isdevised, it is found useful to provide velocityflow cooling by lubricanton a basis of constant flow plus intermittent or auxiliary augmentedflow during the actual work intervals. In this way, a fairly stabletemperature level is maintained for the fin'otion elements required toabsorb the braking energy, and when the bands or discs are actuatedeither for steering or for vehicle braking the cooling system providedis equipped with controls which introduce the supplementary jet flowover the engaging fniction faces.

For this purpose each of the housings of the brakes 46 and 50 forvehicle braking, are equipped with nozzles or jets such as shown in FIG.5 at 309 and 310, one for constant flow cooling, and the other forsupplementary cooling under friction drag load. In the case of thealternate showing of the disc brake 102, 103 of FIG. 8, a pressure feedlead 91 in section 104 is shown to deliver a velocity stream oflubricating oil as a coolant to the splined portion of the of 28,thelead 91 being connected as shown in H6. 14. A similar-lead 91..Wou1d,of course, supply similar coolant passage to the disc brake on theopposite side of the vehicle. I

It is of even greater importance to feed a stream of coolant oil to thesteering brakes 61, 62 and 61', 62' of FIGS. 1 and 2, since these willbe operating under variable torque capacity conditions for aconsiderable percentage of the openatingtime. f

' For this purpose, jet pairs 317' and 318,319 and 321 are shown in FIG..2, and are connected to passages for both constant flow cooling andaugmented cooling, the pair of jets 317 and 319 one for each steeringbrake receiving flow from passage 320 connected to passage 252 of FIG.14, and the pair of jets 318 and .321 receiving fiow'from passages 321and 322 of FIG. 14, which flow is directed from valves 170, 171 topassages 172 and 173 of FIGS. 10 and 14.

It seems clear that the brake and clutch cooling system described,provides at'all times a normal cooling flow for obtaininga temperaturelevel condition as well as a controlled augmentedflow available directlywith energy absorption demand. A further method is described inconnection with FIG. '17.

FIG. 15 is a detailed view of the 'main line pressure regulator valveshown in the schematic view of FIG. 14. This valve is calibrated in itsaction by spring 151 adjustable by screw plug '152 threaded in bore 153of the valve body200.

